balanced mechanical seal free sample

Burgmann cartex single cartridge seal is a world famous cartridge mechanical seal, this seals is popular used for pumps, agitator, mixer, and other rotating machines.

cartex signle cartridge seal is a standard mechanical seal, compared to the modular design with many complex parts, Cartex has been specifically engineered with a simple, robust design to achieve higher overall reliability and performance in extreme conditions.

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Since failure of the mechanical seal on the shaft is the number one cause of pump shutting down, it is important to know the standard terms of the world of mechanical seals.

Mechanical seals are comparable to precision instruments. These seals use margins with many decimal places. The mechanical seal life depends on many factors, and it can vary from a few intense minutes to many trouble-free years. In general it can be stated that: the more attention paid to the mechanical seal and related equipment, the longer they will last.

Years ago, packing materials such as stuffing box packing were used for most shaft seals. These types of shaft seals required a fair amount of leakage to keep the packing properly lubricated and cooled. Until 1915 the mechanical seal was invented. This shaft seal managed to keep the fluid in by using two incredibly flat surfaces (one that rotates with the shaft, and one stationary in the housing). Despite the fact that these treads also need a little bit of ‘leakage’ to create a hydrodynamic layer, this is often not noticeable as this liquid evaporates. Most pumps today have mechanical seals. However, because the parts and surfaces are so delicate, it is also the number one cause of pump failure. This requires a better understanding of this type of seal and its application.

A set of (very flat lapped) treads as primary seal: the minimum distance between these treads, which are perpendicular to the shaft, minimizes the leakage. Often two different materials are used as the tread, a harder and a softer material, to prevent the materials from sticking together. One of the treads is often anti-friction corrosion material such as carbon graphite. A relatively hard material such as silicon carbide (SiC) or ceramic alumina is often used for the other tread. However, when processing abrasive substances, two hard surfaces are normally used.A tread is mounted stationary in a house.

Mechanical seals require a fluid to maintain lubrication. The running surfaces are usually lubricated by a very thin layer of liquid (or gas) between the two running surfaces. Lubrication can also come from another fluid other than the product, depending on the seal requirements.

Pusher seals use a secondary seal that moves axially along the shaft or shaft sleeve to maintain pressure on the running surfaces. This allows the seal to compensate for wear and any less accurate shaft alignment. The advantage is that this seal type is the cheapest. The main disadvantage of this configuration is that the secondary seal can gall on the shaft or shaft sleeve, especially when processing abrasive product.

This is the category in which the bellow seals fall. These seals do not use a secondary seal that must be able to move along the shaft or shaft sleeve to maintain contact. The secondary seals do not move with this type of seal under any circumstances, not even during use. The tread wear is compensated for by an elastomer or metal bellows. A disadvantage of this type of seal is the higher cost price of the seal and that a larger seal must be used in a corrosive environment because the material of the bellows is otherwise too thin.

We speak of a balanced seal if the pressure on the running surfaces caused by the pressure in the system is taken into account. It may sound crazy if the goal is to achieve shaft seal, but a mechanical seal must leak! After all, the running surfaces of the primary seal must be lubricated with the pumped product. When the pressure on the product side exceeds approximately 250 psid, the pressure on the treads can increase to such an extent that no liquid film can form between the surfaces. The lack of lubricating film will cause the seals to wear out very quickly.

To overcome this problem, the balanced seal was introduced in 1938. With a balanced mechanical seal, high pressure is taken into account by adjusting the surface of the tread, which distributes the stuffing box pressure over a larger surface. Balanced seals are easy to recognize, there is a step in the shaft sleeve and/or the running surface. Incidentally, this works a bit more complicated with a metal bellow, but the principle remains the same. Mechanical seals can also be designed to balance for overpressure on both sides of the tread assembly.

Cartridge seals consist of a pre-mounted mechanical seal on a shaft sleeve that can be installed as a whole over the shaft or shaft sleeve. Cartridge seals are very easy to install and the chance of short service life due to suboptimal installation is less probable. It should come as no surprise that cartridge seals are a lot more expensive than the previously discussed seals. On the other hand, there are lower maintenance costs. Incidentally, it is not always possible to apply a cartridge seal if, for example, there is no space in the house.

ARMSTRONGAppendix (iii)Mechanical Seals in HVAC pumpsComments on inside-installed seals and outside-installed balanced seals.Armstrong’s typical mechanical pump seal is designed with stainless steel metalparts, Viton or EPDM elastomer (‘O’ rings & bellows), carbon rotating face andceramic stationary seat. The ceramic seat is secured by an ‘O’ ring in somepump designs and flat gaskets in others. Each is well suited to the specificneeds of the pump.Mechanical seals are generally referred to as ‘leak proof’. Mechanical seal facesare very flat (Seal faces can be ground and polished [Lapped] to a flatness of 1.5Argon light bands) so the gap between them isvery small. Consequently, mechanical seals doleak; yet leak so slowly that water evaporatesbefore collecting sufficiently to be visible. Thefaces must be flat enough to prevent discernibleleakage, yet not so flat as to prevent thenecessary liquid film from entering the gapbetween the faces. A film of liquid between sealfaces is essential for cooling and lubrication. Asa seal rotates, the heat of rotation acceleratesevaporation of the liquid, resulting in a seal thatis virtually leak free.There are two sealing requirements for a mechanical seal to ensure that theliquid stays inside the pump housing and doesn’t leak into the environment. Theareas are typically known as the Primary and the Secondary seals.Secondary sealing is accomplished by the sealelastomer (Usually Viton or EPDM) which is fittedtightly around the pump shaft, preventing liquidfrom migrating along the shaft towardsatmosphere.The two ‘faces’ of a mechanical seal make up the‘primary’ seal. One face (Usually Carbon) rotateswith the shaft and is part of the seal head, alongwith the elastomer and spring(s). The other face(Ceramic seat is typical) is stationary, held inplace by an ‘O’ ring or seal gland plate and gaskets, in the pump casting. Thecarbon ring rotates against the ceramic seat resulting in the Primary seal.As the carbon ring wears from friction, the seal springs (Or bellows, dependingupon the seal design) help push the carbon ring forward and keep it tight againstthe ceramic seat to maintain design tolerances and prevent leakage.Page 1 of 2

An end face mechanical seal is a device intended to prevent or minimize leakage from a vessel through the clearance around a rotating shaft entering that vessel. Perhaps the most common example is the end face mechanical seal used in the water pump of an automobile engine.  Most pumps used in petroleum refineries, chemical plant and pipelines also use end face mechanical seals.

The simplest possible mechanical end face seal consists of a shoulder on a rotating shaft which rubs against a stationary case. This concept is shown in Figure 3.

As with packing or any bearing material, lubrication and cooling are required to prevent heat buildup and wear. Hydraulic pressure tends to force fluid between the faces and provide a lubricating film but the face separation must be kept very small to minimize leakage.  Cooling is provided by the surrounding liquid.  The conceptual design shown in Figure 3 is very simple, but it demonstrates the basic principle of the end face mechanical seals.  Of course, it has functional drawbacks which must be addressed.

Seal face leakage is governed by many variables, but the dominant variable is face separation. A variation of a few micro-inches (millionths of an inch) in the face separation can cause significant changes in leakage.  Unfortunately, shaft movement can amount to several thousandths of an inch.

A practical approach to overcoming shaft movement is to mount one of the seal faces in a flexible manner so that it can move axially. Obviously a desirable feature, this flexibility is a prerequisite for effective seal design.

Figure 4 shows an improved seal as compared to Figure 3. In Figure 4,  The sealing shoulder on the shaft has been removed and replaced with a component which is not permanently attached to the shaft.  This component is called the primary ring.  The “face” of the primary ring rubs against the mating ring.  Since the primary ring and the shaft are two separate parts, an additional sealing device must be used to prevent leakage between the shaft and primary ring.  The flexibly mounted primary ring can compensate for the small variations in movement on the axial plane.  It can also adjust for seal face wear.  Figure 4 is a very simple mechanical seal but it illustrates the concept used by the majority of mechanical seals.  Of course, some additional components are required to preload the faces, transmit torque and provide ease of installation.

Figure 5 shows a more complete mechanical seal including a replaceable mating ring, “O-ring” gaskets, springs, setscrews and various other hardware. As will be seen later, the design of these components may vary considerably according to the required service for the seal.  In addition, the assembled components themselves may be arranged and oriented in various ways to accomplish varying degrees of sealing, reliability and redundancy.

In a mechanical seal, the primary leakage path is between the seal faces.  Naturally, increasing the face separation increases the leakage. In fact, as will be shown later, doubling the face separation can increase the leakage rate by a factor of eight!  This relationship between leakage and face separation provides a powerful incentive for minimizing face separation.  In modern mechanical seals, face separations are so small (on the order of a few microns) that the leakage rate is affected by the surface roughness.  The effective face separation is a combination of the surface roughness of the two mating parts and the fluid film thickness. This concept is illustrated in Figure 6.  The seal manufacturer can control the initial surface finish by lapping and polishing.  A typical seal face is flat to within 23 millionths of an inch.  This degree of flatness is so small that refracted light rays must be used to measure it.  The fluid film is established during initial start-up of the seal by hydraulic forces.

The principle of establishing a fluid film is essential to all seal designs. Most mechanical seals are designed to operate in liquids; these seals require a liquid film.  Designing a seal to operate on a gas film is much more complex.  Whether the film is gas or liquid, it reduces the gross contact between the rotating component and the stationary component.  It also provides lubrication to reduce friction and wear.  Without a stable fluid film, gross rubbing contact could damage the faces.

Mechanical seals may be classified by their design features and the arrangement of those features. The Design category includes the details and features incorporated into a single primary ring/mating ring pair. The Arrangement category includes the orientation and combination of the primary ring/mating ring pair.  Figure 7 illustrates the classification of mechanical seals.

The Design classification considers the details which enter into the features of these components. Some examples of these features are balance, face treatment, rotating element, springs, secondary sealing elements and drive mechanism.  In general, these design features are not completely independent; that is, emphasis of a particular feature may also influence other features.  For example,  selection of a particular secondary sealing element may influence the shape of the primary ring.

By definition, the primary ring is the flexible member of the mechanical seal. The design of the primary ring must allow for minimizing distortion and maximizing heat transfer while considering the secondary sealing element, drive mechanism, spring and ease of assembly.  Many primary rings contain the seal face diameters, although this is not a requirement of the primary ring.  The primary ring always contains the balance diameter.

Balance.   The term “balance” is frequently referred to as the relationship of hydraulic forces on a seal but it is actually a geometric ratio.  Balance ratio is defined as the ratio of the hydraulic closing area to the hydraulic opening area. This ratio is customarily expressed in a percentage.

Figure 8 illustrates the concept of balance. In a seal, hydraulic pressure acts on the back of the primary ring; the resulting force pushes the faces together. This force is called the closing force and this area the closing area.  Similarly, any pressure between the seal faces creates an opening force which tends to separate the faces.  Therefore, the face area is also called the opening area.  The balance ratio is simply the ratio of the closing area to the opening area.

As shown in Figure 8, the area above the seal face outside diameter is disregarded when the closing area is computed. This area is not considered because the pressure is the same all around it;  consequently the contribution of the resultant of the hydraulic forces on this area is zero.

When the closing area is reduced, the closing force is reduced proportionally; this feature can be used to advantage when designing a seal. However, for a seal shape such as shown in Figure 8, the closing area will always be greater than the opening area. In order to make the closing area less than the opening area, the shape can be changed as shown in Figure 9.

The seal shown in Figure 8 is said to be an “unbalanced” seal. Its balance ratio is greater than 100% because of the necessary clearance underneath the mating ring.  Typical balance ratios for unbalanced seals range from 120 to 150%.  The seal shown in Figure 9 is said to be a “balanced” seal; its balance ratio is less than 100%.  The balance ratio of “balanced” seals is typically from 65% to 90%.

The distinction between “balanced” seals and “unbalanced” seals is simply that balanced seals have a balance ratio less than 100%.  A seal with 99% balance ratio is balanced, a seal with 101% balance ratio is unbalanced.

For a given pressure, balanced seals have less face load than unbalanced seals. Therefore, balanced seals are normally used in higher pressures than unbalanced seals.

Primary ring shape.The shape of the primary ring may vary considerably according to the incorporation of various design features.  In fact, the shape of the primary ring is often the most distinct identifying characteristic of a seal.  Figure 10 shows four examples of typical primary ring shapes.

Figure 10a represents a primary ring associated with elastomeric bellows seals. This primary ring has been optimized to take advantage of the elastomeric bellows and a large, single spring.

Figure 10b represents a primary ring with an inserted seal face. Insert faces must be designed with care because temperature differentials can cause differential expansion between the adaptor and the primary ring.  Insert designs can also have problems associated with mechanical stress and distortion.

Figure 10c and 10d show how the shape of the primary ring is influenced by the secondary sealing element. Figure 10c is a primary ring designed to work with a wedge.  Figure 10d is designed to work with an O-ring.  Also, Figure 10c shows an unbalanced shape while Figure 10d is a balanced shape.

Face Treatment.The most common seal face design is a plain, flat surface but there are many special treatments designed for specific applications.  Figure 11 shows some of the more common face treatments.  The plain, flat face is most common.  In general, face treatments are a means of modifying the pressure distribution between the seal faces.  The most common objective is to increase the opening force and thereby reduce the magnitude of the mechanical contact.  Face treatments may be considered to produce hydrostatic or hydrodynamic forces.

Rotating Element. Although most illustrations have shown the primary ring to be rotating with the shaft, either the primary ring or the mating ring may be used as the rotating element.  Seals with rotating primary rings are said to be “rotating” seals; seals with stationary primary rings are said to be “stationary” seals.  Because the springs are always associated with the primary rings, sometimes the distinction is made as “rotating springs” versus “stationary springs”.

For convenience, rotating seals are used in most equipment. Pump shafts are already made of a comparatively high grade material and manufactured to close tolerances. This makes pump shafts well suited for rotating seal applications.   Assembly of rotating seals can generally be done directly on the shaft or by using a relatively simple sleeves.  Figures 5, 8 and 9 all show rotating seals with stationary mating rings.

Stationary seals have some advantages over rotating seals. In small, mass produced seals for modest services, the entire seal may be placed in a package which minimizes shaft and housing requirements for the equipment.  Figure 13 shows a low cost stationary seal.   Stationary seals are also used to advantage in large sizes or at high rotational speeds.  Above 5,000 to 6,000 fpm, a rotating primary ring (which is flexible, by definition) may require dynamic balancing (for rotational imbalance) in order to operate in a stable mode.  A stationary primary ring does not require this balancing.  On the other hand, the stationary seal does require a close bore tolerance.  This close bore tolerance is usually a second manufacturing operation on most equipment.  Stationary seals sometimes incorporate special design features such as auxiliary liners, sleeves or adapters to help retain the mechanical advantages of the seal.  Figure 14 shows a stationary metal bellows seal used in high temperature centrifugal pumps.

By definition, the mating ring is the non-flexible member of the mechanical seal. The design of the mating ring must allow for minimizing distortion and maximizing heat transfer while considering ease of assembly and the static secondary sealing element.  The mating ring can contain the seal face diameters, although this is not a requirement of the mating ring.  To minimize primary ring motion, the mating ring must be mounted solidly and should form a perpendicular plane for the primary ring to run against.  Figure 15 shows some typical mating rings.

Floating. Figure 15d is the floating ring type.  This design offers the flexibility of using teflon, grafoil, or an O-ring (Figure 15c) for the secondary sealing item.  If teflon or grafoil is used, a pin should be added to prevent rotation due to the low coefficient of friction of these materials.

The secondary sealing elements are gaskets which provide sealing between the primary ring and shaft (or housing) and the mating ring and shaft (or housing). They are called secondary sealing elements because their leakage path should be secondary to the seal face leakage.  Loading by hydraulics or mechanical force makes the secondary seal tight in its confined area.  The secondary sealing element for the mating ring is always static axially (although it may be rotating).  Secondary sealing elements for the primary ring are described as being either pusher or non-pusher in the axial direction. The term pusher is applied to secondary seals that must be pushed back and forth by the movement of the shaft or primary ring.  A non-pusher secondary seal is a static seal for the primary ring.

Pusher type secondary seals have the disadvantage of damaging the surface to which they must seal. This damage, called fretting, is caused by the cyclic movement of the secondary sealing element as the shaft rotates.  In contrast to the pusher design, a non-pusher secondary sealing element could not cause any fretting.

On the other hand, this rubbing and dragging effect is also an advantage of the pusher design because it adds damping and therefore stability to the seal. This damping can make a significant difference in performance for some services.

Figure 16 shows examples of pusher, non-pusher and static secondary seals. The pusher design may use O-rings, wedges, etc.  The non-pusher is always some sort of bellows with a static section.  Mating rings use various static gasketing and O-ring designs.

Bellows. Figure 16a is a full convolution elastomeric bellows.  It offers the greatest possible flexibility to the front section of the primary ring.  The front section of this bellows has minimum contact with the shaft or sleeve, thus minimizing wear and hang-up.  The large tail section provides a considerable sealing area to compensate for imperfections in the shaft.  There are also are variations of the full convolution design  that use a half convolution.   Its ability to accept axial motion is reduced, due to the one-half convolution design.

Figure 16b is a bellows made out of teflon or glass-filled teflon combinations. This seal which is designed for the extremes of corrosive environments, provides the advantages of the non- contacting bellows convolution.  Because it is made out of teflon, support rings or a drive collar is required to attach the tail section or the bellows to the shaft.  Due to the requirements of flexibility, the convolutions must be considerably larger in cross-section than the typical elastomeric bellows design.

Figure 16c is an all-metal bellows style seal. It has the inherent advantage of flexibility associated with bellows design for high temperature applications.  Due to its all-metal construction, it offers considerable freedom of design since it is not restricted to the temperature and chemical limits of elastomers.  Metal bellows are constructed of individually welded metal leaves approximately .004/.012 thick.  A large number of leaves are required to provide the maximum amount of flexibility associated with other styles of bellows seals.  The mechanical closing force that is provided by the spring on other seal designs is accomplished by stressing the bellows from free height in this design.  Metal bellows seals are inherently balanced because of the manner in which the bellows becomes distorted when pressurized.

“V” Rings, “U” Cups and Wedges. Figure 16d is a wedge. Wedges are typically made of TFE material. They are considerably more flexible than the “V” ring or U-cup arrangement because they operate on the ball-and socket principal. Special manufacturing fits are not a requirement for wedges because of the shallow angle of contact between the primary ring and the shaft.  However, it does require a polished surface for effective sealing (32 rms with polish).

In addition to wedges, there are U cups and V rings.  Their  construction as a secondary seal offers limited amounts of flexibility and requires extremely close tolerances on the cross-section fits.  The “V” rings are generally manufactured in TFE or TFE-filled material, requiring pre-loading of the rings to activate the point contact on the lips.  Because of the “V” ring design, it can be used at high pressures but it is the least flexible, and most rigid of the secondary seal designs. Because of it’s construction, the “V” ring cannot flex to compensate for motion and it must be pushed along the shaft to take up wear at the seal faces.  Shaft and primary ring finish must be highly polished (15 rms) in order for this type of secondary seal to operate leak-free.

O-Rings. Figure 16e is the O-ring.  This is by far the simplest and most popular secondary sealing element.  It has been used successfully over a wide temperature range and in a variety of fluids.  O-rings are considered to be self-energizing seals and do not require much mechanical preloading.  This feature allows O-rings to be used at very high pressures.  O-rings are offered in a complete range of chemical resistant and general service compounds.  Buna-N, Neoprene, Ethylene-Propylene, Viton, and Kalrez are typical materials selected for a variety of service conditions.  On the other hand, Teflon does not make a good O-ring material, particularly if the O-ring is to be dynamic.

There are encapsulated O-rings. This is an attempt to obtain the chemical resistance offered by Teflon and the flexibility of the elastomeric O-ring.  Unfortunately, the result is dominated by the hardness of the Teflon outer shell.  The recommended surface finish of the shaft/sleeve is 15 rms.  The encapsulated “O” ring is sensitive to temperature fluctuations.  Because of these limitations, encapsulated O-rings are usually recommended only for static sealing.

In every mechanical seal there is always a need for keeping the faces closed in the absence of hydraulic pressure.   Generally, a mechanical device in some form of spring is used.  Figure 17 shows some of the different types of springs used in mechanical seals.

Single spring.A single spring seal has the advantage of comparatively heavy cross section coil which can withstand a higher degree of corrosion.  Another advantage is that single springs do not get clogged by viscous liquids.  The disadvantage of a single spring is that it does provide uniform loading characteristics for the faces.  Also, centrifugal forces may tend to unwind the coils.  Single springs also tend to require more axial space and a specific spring size is required for each seal size.

Multiple springs.Multiple springs are usually smaller than single springs and provide a more uniform load at the faces.  The same spring size can be used with many seal sizes by simply changing the number of springs that are used. Multiple springs resist unwinding from centrifugal force to a much higher degree than a single coil spring since the forces act differently.  The most obvious disadvantage of small springs is the small cross section wire.  This makes the smaller springs subject to corrosion and clogging.

Wave spring. The next form of spring generally considered is the wave type, simple described as a washer into which waves have been formed to provide a given amount of mechanical loading . The main reason for using this type of spring is that it requires even less axial space than the multiple spring design.  On the other hand, special tooling must be made for best manufacturing results.  Further, the tempering required on this design limits materials to those which are not as corrosion resistant as the high grade stainless and Hastelloy groups.  Also, when using wave springs, a greater change in loading for a given deflection must be tolerated.  That is, a great deal of force loss or force gain, with comparatively small axial movement must be expected.

Metal bellows.  A metal bellows is actually a combination of a spring and secondary sealing element.  Welded edge metal bellows resemble a series of Belleville washers.  Formed bellows may be used to reduce the quantity of welding; however, a formed bellows has a much higher spring rate than a welded bellows.  The bellows thickness must be selected for resistance to pressure without an excessive spring rate.  The welding technique and bellows shape must be selected for maximum fatigue life.

The term “hardware” is used to describe the various devices which hold the other components together in the desired relationship.  For example, a retainer might be used to package the primary ring, secondary sealing element and springs into a single unit.  Another example of hardware is the drive mechanism which is necessary to prevent axial and rotational slippage of the seal on the shaft.

A drive mechanism is required because of the torque created between the seal faces. Both static and dynamic drives are required. The static drive is only required to hold an axial position and transmit torque.  The dynamic drive must transmit torque and allow for the axial flexibility of the primary ring.  Figure 18 shows some variations of drive mechanisms.  Dedicated devices made of strong materials are said to be positive drives and are generally preferred in the heavy duty seals.

Key drive.This is one of the more rugged forms of drive. In high pressure, comparatively large size units, the ruggedness of this type of drive is in keeping with the balance of the sturdy features incorporated in seals for high pressure applications.

Elastomer Drive.This is not a positive drive method and is generally limited to light duty seals; however, it is extremely simple, economical and effective for some services.

Spring drive.Another effective drive mechanism for light duty seals is to incorporate the drive into the springs.  Care must be taken in this design as to the direction of rotation, corrosion rate and spring rate.

Although all end face mechanical seals must contain the five elements described above, those functional elements may be arranged or oriented in many different ways. Several dimensional and functional standards exist, such as API Standard 682 – Shaft Sealing Systems for Centrifugal and Rotary Pumps, which describes the configurations for used in Oil & Gas applications. Even though the scope of API 682 is somewhat limited, it may be extended to describe end face mechanical seals in general.

Configuration refers to the number and orientation of the components in the end face mechanical seal assembly. For example, springs may be rotating or stationary.  Single or multiple pairs of sealing faces may be used.  For multiple seals, the individual pairs of sealing faces may be similarly oriented or opposed.  Containment devices such as bushings may or may not be used as part of the configuration.

Single seals.  The vast majority of all seals fall into the single seal category. In this category the seal can be mounted so that it is inside the process liquid or outside the process liquid.  Inside mounted seals are easier to cool and generally can seal higher pressures.  Also the direction of leakage is opposed by centrifugal force.  Inside mounted is by far the most popular, see Figure 19.

An outside seal is shown in Figure 20.  Outside seals have minimal contact with the process liquid.  This is an advantage in sealing corrosive liquids providing that overheating is not a problem.  A disadvantage of outside seals is that the direction of leakage is the same as centrifugal force.

Multiple seals.  Multiple seal arrangements can provide environmental controls and/or redundancy.  The most common types of multiple seal arrangements were previously called double and tandem but are called Arrangement 2 and Arrangement 3 by API 682.  The double seal emphasizes environmental controls while the tandem seal emphasizes redundancy.

The principal purpose of the tandem seal is to provide redundancy. The two seals are oriented so that the outer seal can accomplish the sealing task if the inner seal fails. A tandem arrangement is shown in Figure 21.  Because of this redundancy, the outer seal is sometimes called the “backup” or “safety” seal.  Tandem seals use a buffer fluid to lubricate and cool the outer seal.  The inner seal operates in the process liquid and is cooled by that liquid.  Any leakage from the inner seal must be contained by the outer seal and buffer system.  The buffer fluid is normally at a pressure less than the stuffing box pressure.  In order to keep the buffer pressure low, the buffer system is vented continuously.

Tandem seals (API Arrangement 2) are also used in “stages” (perhaps this is the origin of the name “tandem”)  when process pressures are extremely high. When tandems are used in stages, the buffer system is pressurized to some intermediate pressure, typically half the stuffing box pressure.

The double seal (API Arrangement 3) uses two primary ring/mating ring pairs oriented so that a pressurized barrier liquid is maintained between them as shownin Figure 22. On double seals, the inner seal seals between the process liquid and the barrier liquid.  Because the barrier system is at a greater pressure, any leakage is barrier liquid.  The outer seal seals between the barrier liquid and the atmosphere.  In most cases the barrier liquid is circulated and cooled to prevent heat buildup.

Notice that Figure 21 and 22 are identical.   The only differences are the design of the faces and balance diameter and these details are not shown.  This is somewhat controversial and many would describe both Figure 21 and 22 as being “tandem” because of the orientation of the components.    Traditionally, a “double” seal would have the primary rings in a back-to-back orientation; however, for purposes of operation, either face-to-back (shown) or back-to-back may be used.

The biggest problem with the double seals is maintaining the barrier liquid at a higher pressure than the stuffing box. It is generally recommended that a 20 psi or 10% differential in favor of the barrier be maintained at all times.

In the past, double seals were described as primary ring/mating ring pairs which faced in opposite directions.   A classic double seal is shown in Figure 24.  This is usually true but is not a requirement of the double seal arrangement.  Similarly, a tandem arrangement was frequently described as seal pairs facing the same direction, see Figure 21.  Again, this is not a requirement.  In a tandem arrangement, the buffer fluid is at the lower pressure and is continuously contaminated by leakage of the product across the primary seal.  In a double seal arrangement, the barrier fluid is at the higher pressure and the product fluid is continuously contaminated by leakage of the barrier fluid across the primary seal.

Double seals (API Arrangement 3) are used with API Piping Plans 53A, 53B, 53C or 54.   Tandem seals (API Arrangement 2)  are used with API Piping Plans 52and 55.

The term “adaptive hardware” is applied to hardware designed to simplify the incorporation of the primary ring and mating ring into the equipment which requires the mechanical seal. The most common examples of adaptive hardware are the sleeve and gland.

When the components are pre-assembled onto a sleeve and gland plate, the complete assembly is called a cartridge seal. This complete assembly can be easily slid onto the shaft and bolted in place thus reducing the potential for installation errors. Some cartridge seals use regular component seal parts whereas other cartridge seals might use specific purpose parts. API 682 specifies that only cartridge seals are acceptable to the standard.

Figure 23a shows the plain integral type gland plate which, if lapped on the sealing surface and gasketed perfectly to the face of the box without distortion, can serve as a sealing face against the rotating seal head. It is an item utilized on many compressor installations, showing that when applied correctly, even a gas such as Freon can adequately be sealed with this type of design.  It becomes a desirable feature, especially where axial room is at a premium.  It must be considered only where practical.  If a relatively large bolt circle in relation to the shaft size is inherent in the unit, this design may prove uneconomical.

Figure 23b is a plain end plate with a pressed-in seat element. Some seal manufacturers advocate this design and it can be utilized where its stresses, replacement features and economics can be tolerated.

Still another auxiliary feature is termed the quench type gland plate. This gland type incorporates an entrance for liquid behind the seal face on the atmosphere side.  Two other desirable features are to be gained from this quench gland-one is the safety provided by the throttling effect on the escaping liquid if the seal fails; the other relates to toxic liquids.  If the liquid handled is toxic-instead of using the quench gland with a secondary liquid, the liquid can be vented to a safe area.

Sleeves are employed as an adapter from the seal requirements to the shaft requirements of the equipment. For example, the seal may require different diameters, as in a balanced seal.  Figure 24 shows the two most common types of shaft sleeve.

Figure 24b is the “cartridge” or “package” sleeve.  This design allows the complete seal, gland and sleeve to be assembled before it is installed in the pump.

is in a DGS. Instead, the highest pressure is between the seal faces, and the external pressure will flow into the atmosphere or vent into one side and into the compressor from the other side. This increases reliability by keeping the process out of the gap. In pumps this may not be an advantage as it can be undesirable to force a compressible gas into a pump. Compressible gases inside of pumps can cause cavitation or air hammer issues. It would be interesting, though, to have a non-contacting or friction-free seal for pumps without the disadvantage of gas flow into the pump process. Could it be possible to have an externally pressurized gas bearing with zero flow?

There is another advantage in compressors in that there is no flow across the face as there is in a DGS. Instead, the highest pressure is between the seal faces, and the external pressure will flow into the atmosphere or vent into one side and into the compressor from the other side. This increases reliability by keeping the process out of the gap. In pumps this may not be an advantage as it can be undesirable to force a compressible gas into a pump. Compressible gases inside of pumps can cause cavitation or air hammer issues. It would be interesting, though, to have a non-contacting or friction-free seal for pumps without the disadvantage of gas flow into the pump process. Could it be possible to have an externally pressurized gas bearing with zero flow?

All externally pressurized bearings have some sort of compensation. Compensation is a form of restriction that holds pressure back in reserve. The most common form of compensation is the use of orifices, but there are also groove, step and porous compensation techniques. Compensation enables bearings or seal faces to run close together without touching, because the closer they get, the higher the gas pressure between them gets, repelling the faces apart.

54MB? 14 CO/V/(Z //V hwy United States Patent This invention relates to a mechanical seal for sealing a shaft with respect to an enclosure from which it extends, such as the stufling box of a centrifugal pump. More particularly, this invention relates to a seal of the so-called balanced type wherein springs are utilized for creating a portion of the force urging the sealing faces together and for compressing the shaft packing and in which all springs are fully protected from contact with the fluid within the enclosure from which said shaft extends.

Mechanical shaft seals have been in wide commercial use for at least thirty years and have been constructed in a wide variety of designs. Throughout this period, one of the major problems with certain types of mechanical seals has been to protect the springs therein from corrosion by the fluid within the enclosure being sealed. This problem has been dealt with in a variety of ways but it has not previously, to my knowledge, been solved in a completely satisfactory fashion.

Referring now, for illustrative purposes, as aforesaid, to the problem of providing a seal around the shaft of a centrifugal pump, a common type of mechanical seal utilizes a stationary, annular, sealing face affixed to the pump housing and a rotating, annular, sealing face aflixed to the pump shaft. The two sealing faces are held in sealing engagement with each other and hence provide an effective seal against the passage of fluid from one side of the pump housing to the other along the shaft. Various types of resilient means, commonly metallic springs, are provided for initially urging and holding the sealing faces in sealing engagement with each other and, in many cases, for holding the shaft packing in sealing engagement with the shaft. Where the fluid either liquid or gaseous, being handled by the pump is corrosive the springs must be protected from contact with the fluid or else the seal will deteriorate rapidly. In low pressure applications, this has been accomplished by what is known in the trade as an outside seal, that is, a seal within the rotating portion thereof, including the springs, is placed outside of the housing and the mutually contacting sealing faces. This arrangement effectively protects the springs from the fluid within the pump but thus far it has been limited to relatively low pumping pressures, such as 40 or 50 p.s.i. For higher pressure applications, it has been found essential to place the rotating portion of the seal in communication with the interior of the pump housing in order to utilize the pressure of the fluid being pumped to assist in urging the sealing faces together. This is effective for sealing against high pressures, but where the fluid being pumped is of corrosive nature, this construction exposes the rotating parts of the seal, including the springs to the corrosive effect of such fluid. This can usually be handled with respect to all parts of the rotating portion of the seal except 3,l8,53. Patented Aug. 3, 1$65 the springs by selection of suitable corrosion-resistant materials. As to the springs, no materials have yet been found from which satisfactory springs can "be made and which also can resist corrosion by some of the fluids with which the sea-ls are"required to be used."

.Some attempts have been made to meet this problem by arranging flexible housings or jackets around the springs in the rotating portion of the seal but these have not been particularly satisfactory because either they have often leaked and hence failed to protect the springs anyway, of

Other attempts to meet this problem have involved the use of a so-called double seal wherein a pair of seals are arranged in a somewhat backato-back relationship with respect to each other and are located at opposite ends of a stuffing box with the space between said seal-s being filled with a pressurized sealing fluid, such as oil. This has effectively protected the springs against attack by the fluid being pumped but it is expensive, introduces substantial operating and maintenance problems, often requires more space than is conveniently"available and sometimes results in the sealing fluid leaking through the inside one of the pair of seals and contaminating the fluid being handled by the pump.

Further attempts to meet this problem have involved a type of .seal which is basically an outside seal, in that the rotating portion of the seal is outside of the housing and the sealing faces, but wherein a portion of the fluid being pumped is permitted to pass axially along the shaft to a zone within the rotating portion of the seal itself so that the fluid exerts a pressure which assists in holding the sealing faces together in a manner generally similar to that accomplished in the above-mentioned inside seals. In such a seal, often called a balanced seal, the springs by which the sealing faces are initially urged together are outside of the equipment being sealed and, hence, partake of the above-mentioned advantages of an outside seal. However, since the fluid in this case is introduced into a zone within the rotating portion of the seal, this introduces the further problem of effecting a seal between said rotating portion and the shaft. This problem is seriously complicated by the fact that the rotating sealing member is necessarily free to float slightly on the shaft, both axially and angularly, in a manner to follow slight variations from exact perpendicularity of the sealing faces with respect to the axis of the shaft. Ordinary shaft packings between the rotating sealing member and the shaft, even though of a resilient nature, have not proven satisfactory. Some attempts have been made to use spring-backed shaft packings for this purpose but all such attempts, insofar as I am aware, have resulted in these springs being exposed to the fluid being handled by the pumps and, accordingly, subject to deterioration.

("1) To provide means for mechanically sealing a rotary shaft with respect to a wall through which it extends, which seal is of the outside balanced type and has springcompressed shaft packing and wherein the springs are fully protected from contact with the fluid being handled by the equipment of which said shaft is a part.

(2) To provide a mechanical seal for equipment, as aforesaid, in which pressure fluid will be introduced into the interior of the rotating portion of the seal for pressure balancing purposes and in which the rotating sealing member will be permitted to float both axially and angularly with respect to the axis of the shaft but wherein said rotating sealing member will be effectively sealed with respect to the shaft.-

(3) To provide a mechanical seal for equipment, as aforesaid, in which shaft packing is provided between the rotating sealing member and the shaft and is backed by resilient means, said resilient means being protected from contact by the fluid being handled by said equipment.

(4) To provide a mechanical seal for equipment, as aforesaid, which will be compatible with previously known designs for such seals and hence is adapted to a variety of uses.

(6) To provide a mechanical seal, as aforesaid, in which the resilient means by which said shaft packing is compressed into position, may be readily modified or interchanged to fit a variety of different operating purposesand circumstances.

FIGURE 1 is a central longitudinal sectional view taken through a mechanical seal constructed according to the invention and shown in association with a shaft and a housing.

General description In general the invention comprises a mechanical seal of generally the outside type of construction, which seal has a zone at its radially inner side for receiving pressure fluid from within the equipment on which the seal is used. The pressure fluid supplies a balancing pressure on the axially outerside of the rotating sealing member which urges said rotating sealing member into sealing engagement with the nonrotating sealing member. Shaft packing is arranged between the rotating sealing member and the shaft on which it is mounted at a point axially outwardly. of the afore-mentioned zone. A spreader ring of suitably corrosive-resistant material is arrangedvbetween the aforementioned zone and the shaft packing. Rods are anchored in the spreader ring and extend through the shaft packing away from the zone. The rods are engaged by resilient means on the side of the shaft packing opposite the zone and suitable tension is applied thereto for pulling the spreader ring against the shaft packing and holding said shaft packing under suitable pressure so that it sealingly engages the shaft and the rotating sealing member. The resilient means, which may be coil springs surrounding the rods, are thereby protected from contact with the fluid but insure efiective operation of the shaft packing.

Referring now particularly to FIGURE 1, there is indicated a housing 1 which will be understood to refer to the housing of the pump, mixer, autoclave or other equipment with which the seal is being used. A shaft 2 extends through an opening 3 within said housing, said shaft being supported for rotation with respect to the housing All.

by an appropriate bearing means (not shown). A nonrotating seal member 4, commonly called an insert in the trade, encircles said shaft and is arranged within a suitable mounting ring 6. The ring 6 also encircles said shaft and is fastened to the housing 1 by any convenient means, such as screws,"one of which is indicated at 7. While the insert can be made of a variety of materials, as is well known to those skilled in the art, frequently it is made of a ceramic of various types and hardnesses depending on particular use requirements. The insert 4 has a planar axially outwardly facing surface 5 which functions as a stationary sealing face. The mounting ring 6 is made from a metal or allow suitable physically and chemically for the ambient conditions.

A rotating sealing member 11 encircles the shaft and has a radially extending body portion 13 whose radially inner end is spaced slightly. from said shaft to provide a passageway 12 therebetween. The sealing member 11 also has a sleeve portion 14 which extends axially away from the insert 4 and which surrounds and is spaced a substantial radial distance from the shaft 2. The member 11 also has an axially projecting annular flange 15 whose end surface 16 engages the surface 5 of the insert 4 and thus functions as the rotatable sealing face of the mechanical seal. The internal wall of the sleeve 14 defines an annular chamber 17 around the shaft 2.

A collar 18 is affixed to the shaft 2 by any convenient means, such as by set screws 19. A plurality of springs 20 are mounted in circumferentially spaced recesses 21 in the collar 18 and extend axially therefrom and engage the outer axial end of sleeve 1 The springs 20 urge the rotating sealing member 11 inwardly (rightwardly) so that surface 16 sealingly engages surface 5 to thereby prevent escape therebetween of fluid from within the housing 1.

The collar 18 has an annular boss 22 which projects a small axial distance into the chamber 17. A spreader ring 23 having, in the present case where it is used with packing of substantially V-shape in cross section, a tapered portion 24 is disposed in the chamber 17 encircling the shaft 2 and said ring is spaced axially from boss22. A plurality, here two, of substantially V-shaped, annular packing elements 26 and 26a are disposed in the chamber 17 and encircle the shaft 2 between the boss 22 and the spreader ring 23. The tapered portion 24 fits between the lips of the adjacent packing element 26aand the tapered central portion of said element 26a fits between the lips of the other element 26. Thus, when spreader ring 23 is urged toward boss 22, the lips of the packing elements 26 and 26a are spread apart into sealing engagement with the shaft 2 and the internal wall of sleeve 14.

- spring 32 encircles each rod 27 and seats at one of its ends against the head 31 and seats at its other end against the inner end wall of recess 29. The springs 32 are under compression so that the rods 27 and thereby the ring 23 are continuously urged axially outwardly with respect to the housing 1 to thereby continuously urge the lips of the packing elements 26 and 26a apart and thereby into sealing engagement with the shaft 2 and the internal wall of sleeve 14, as above described.

The space between the spreader ring 23 and the body portion 13 defines a pressure zone 36 which continuously communicates through passageway 12 with the interior of housing 1. Thus, the pressure fluid in said housing can travel to zone 36 and act therein to urge the sealing mem ber 11 inwardly to cause surface 16 to sealingly engage with surface 5. Such pressure also urges the spreader ring 23 axially outwardly to maintain the packing eleavailable to maintain such seal.

If desired, a shroud 37 may encircle the collar 13 and the sealing member ll. The shroud 37 is secured to collar ad by screws 33, said screws extending through axially elongated openings 39 (FEGURE 3) in the shroud so that the shroud 37 and the collar may move axially with respect to each other a limited distance. The shroud 37 also has openings 41 whereby the set screws 19 may be manipulated. This arrangement holds the collar 18 and sealing member 11 in association with each other so that they can be mounted on the shaft 2 as a unit. However, it permits relative axial movement of the collar 18 and the sealing member 11 as necessary to establish the proper sealing relationship between surfaces 5 and 1.6.

The remainder of the seal operates in a normal manner but same will be mentioned briefly to ensure a complete understanding of the invention. Rotation of the shaft 2 acts through the collar 18 and the drive pins 343 (or the shroud 37 if no drive pins are used) to rotate the sealing member 11. The springs 2% effect an initial engagement between the rotating sealing member 11 and the nonrotating sealing member 4 while the pressure builds up within the housing 1. When such pressure does become available, the pressure fluid enters through the passageway 12 into the pressure zone 36 and acts against the sealing member 11 to urge same in an axially inward direction and thereby hold same snugly against the nonrotating sealing member. This effect continues as the pressure builds up within the said housing and will in a well-known manner maintain the nonrotating and the rotating sealing faces snugly together in sealing relationship with each other. Pressure as high as 490 psi. have been successfully sealed by seals made according to the invention.

The sealing member 11, the ring 23 and the packing elements as and 26a, which are the only parts of the mechanical seal which contact the pressure fluid, can be made of any material desired by those skilled in the art, particularly a suitable chemically inert material so that corrosion thereof is not possible. For example, the sealing member 11 is usually made of a carbon of any desired type and hardness and the ring 23 and packing elements 26 and 26a are frequently made essentially of a tetrafluoroethylene resin.

In a mechanical seal for sealing a rotating shaft with respect to a housing through which said shaft extends, wherein a first nonmetallic sealing member is alfixed to said housing, a second nonmetallic sealing member encircles said shaft and is radially spaced therefrom, said second sealing member having a sealing surface engageable with a sealing surface on said first sealing member, said second"sealing member having a sleeve extending away from said first sealing member and surrounding said shaft to define a pressure chamber between said sleeve and said shaft for reception of a balancing pressure fluid from within said housing, and driving means aflixed to said shaft for efliecting rotation of said second sealing member with said shaft, the improvement comprising:

substantially V-shaped shaft packing means within said pressure chamber spaced axially from said wall and adapted to engage sealingly said shaft and said sleeve, said shaft packing means being of such characteristics that an axial pressure thereon will cause same to spread radially inwardly and outwardly;

a plurality of circumferentially spaced rods anchored to said compression ring and extending axially through said shaft packing means and through said passages into said recesses, said rods having enlarged heads, a coil spring disposed within each of said recesses surrounding said rod therein and bearing at one end against said rod head and bearing at the other end against the bottom of said recess for urging said compression ring against said shaft packing means and thereby urging said shaft packing means against said sleeve and said shaft whereby to seal the end of said pressure chamber remote from said wall and thereby prevent escape of pressure fluid therefrom while preventing contact of said springs by said pressure fluid.

From the operational point of view of centrifugal pumps, it becomes essential to correctly align the pump and the drive to ensure the mechanical seal functions properly. Attention shall be given to is nozzle loads. During the design as well as during the actual installation, the consideration of the nozzle loads is important. Higher nozzle loads beyond allowable values could lead to deformed casings and may be detrimental to mechanical seals due to rubbing of the shaft at the clearances. The sizing of the shaft in case of end suction pumps (and also the overhang) has to be controlled, which could result in excessive deflection at the mechanical seal faces.

When it comes to reliability of sealing the process liquid, a dual seal arrangement is the preferred choice. There are three arrangements defined in API 682: arrangement 1, 2, and 3. The arrangement 1 is the single seal arrangement. The arrangement 2 is the dual seal arrangement with unpressurized buffer liquid at the outboard seal. Finally, arrangement 3 is the dual seal arrangement with the pressurized barrier liquid at the outboard seals. With the barrier liquid being pressurized in arrangement 3, there is no leakage of process liquid to the atmosphere, and hence it is the most reliable option when it comes to applicability of stringent environmental norms from the point of view of the end user.

However, in order to ensure proper functioning and reliability of dual seals, the operational environment of the pump, piping, seal support system, and monitoring systems play a vital role. There are typically four API piping plans for seal support systems: API Plan 53 A, B, and C, and Plan 54.

All three variations of Plan 53 are similar from the point of view that they circulate the barrier fluid using the pumping screw inside the mechanical seal, but the methods of pressurizing the barrier fluids are different. Plan 53A uses direct pressurized nitrogen to pressurize ¬fluid in the reservoir. This plan is popularly used in most of the cases due to less complexity and also availability of nitrogen pressurizing source at site. However, to ensure reliability, one has to be careful about the absorption of nitrogen gas into the barrier ¬fluid. The amount of gas being absorbed is proportional to the pressure of the barrier system. The barrier ¬fluid with absorbed gas then reaches the seal faces due to circulation and at the ¬fluid film, due to depressurization, the gas may come out and hamper the seal performance. This is a reliability concern, and hence most of the seals with Plan 53A are limited to 10 bar (gauge) pressure. Plan 53B uses a bladder accumulator as a means of pressurization of barrier fluid. This overcomes the limitation of Plan 53A and the absorption of nitrogen into the barrier liquid, which limits the system pressure, which can be used in high pressure applications. The advantage of the Plan 53B is that it can be used in remote locations where the external source of pressurization is not available. The pressure of barrier liquid is maintained due to the expansion of the bladder inside the accumulator, which also enables the supply of make-up barrier liquid to compensate for a small amount of leakage of barrier -fluid. However, the monitoring of the liquid level in the reservoir is not possible, and as such, the sizing of accumulator considering the seal leakage and maintenance interval is critical. As the bladder expands to compensate for seal leakage, it needs to be refilled with barrier liquid. The usual cycle of refill is 25 to 28 days. Considering this as a basis, the size of the accumulator and the pre-charge pressure of nitrogen is estimated.

Plan 53C uses a piston as a means of pressurization of barrier ¬fluid inside the accumulator. The advantage of this design is that it uses the process fluid pressure from the seal chamber directly on the bottom side of the piston, whereas top side is exposed to the barrier liquid. The pressurization is achieved by the difference in the areas. The area exposed to process liquid is larger and is designed with ratios ranging from 1:1.1 to 1:1.25. As the seal chamber pressure is being used as a reference, the system itself takes care of process pressure fluctuations. However, as the piston is in direct contact with the process fluid, the material selection becomes essential. Also, the properties and quality of process ¬ fluid shall be carefully evaluated, it should not hinder the movement of the piston within the accumulator. Another important factor is the dynamic sealing of the process fluid from the barrier fluid. The failure of the piston seal will result in the equilibrium of pressures on both sides of piston, and because of the piston movement, friction and drag come into play. Thus, the plan is not so reliable for low pressure applications and recommended to be used in the applications with pressures greater than 7 bar (gauge).

Although a mechanical seal is a critical piece of equipment, it shall not be treated in isolation and due consideration should be given to the operating environment of the pump, seal support system, and most importantly, the perfect selection for the given application.

Abhijeet Keer is a design engineer who has been working in the fi eld with centrifugal pumps for over seven years. With strengths in mechanical construction and materials, he has gained valuable knowledge working in design with major players in pump industry, such as KSB Limited and Kirloskar Brothers Limited. He completed his Bachelor’s Degree in Mechanical Engineering from University of Mumbai, India. His professional experience covers new product design and developments, material selection and application engineering, and complete mechanical constructions.